Hydraulic transmission systems

ABSTRACT

A control system in hydraulic transmission apparatus comprising a prime mover, a variable swept volume positive displacement pump driven by said prime mover, a hydraulic machine operated by said pump, said machine being subjected, in use, to variable loading and being required to operate at different rates; the control system comprising, pump control means for varying the swept volume of the pump, prime mover control means for varying the power supplied to the prime mover, one operating element common to both the pump and prime mover control means and arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds a predetermined value dependent on the displacement of the operating element, said pump control means permitting the swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means including a speed responsive governor and being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an element associated with the prime mover control means for effecting reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latter&#39;&#39;s maximum available torque output.

United States Patent [191 Gatiss 4 [45'] Apr. 16, 1974 HYDRAULIC TRANSMISSION SYSTEMS [75] lnventor: Albert Leslie Gatiss, Sutton Coldfield, England [73] Assignee: Brockhouse Engineering Limited,

Stafford, England 22 Filed: Jan. 18, 1973 [21] Appl. No 324,642

[56] References Cited UNITED STATES PATENTS I 1,259,090 3/1918 Ferris et a1 60/19 1,981,805 11/1934 Kacer et a1. 60/19 3,003,309 10/1961 Bowers et a1. 60/431 3,284,999

Primary ExaminerEdgar W. Geo ghegan Attorney, Agent, or Firm-Merriam, MarshalLShapiro & Klose 1 9 ABSTRACT A control system in hydraulic transmission apparatus 11/1966 Leasem, 60/431 comprising a prime mover, a variable swept volume positive displacement pump driven by said prime mover, a hydraulic machine operated by said pump, said machine being subjected, in use, to variable loading and being required to operate at different rates; the control system comprising, pump control means for varying the swept volume of the pump, prime mover control means for varying the power supplied to the prime mover, one operating element common to both the pump and prime mover control means and arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds a predetermined value dependent on the displacement of the operating element, said pump control means permitting the swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means including a speed responsive governor and being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an element associated with the prime mover control means for effecting. reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latters maximum available torque output.

10 Claims, 9 Drawing Figures PATENTEHAPR as 1924 SHEET 6 OF 7 i 1 HYDRAULIC TRANSMISSION SYSTEMS BACKGROUND or THE INVENTION 1. Field of the Invention This invention relates to a control system in or for hydraulic transmission apparatus, which apparatus is of the kind, herein called the kind specified, which comprises a variableswept volume positive displacement pump driven by a primemover, for example, an internal combustion engine, which pump is arranged to operate a hydraulic machine such as a positive displacement motor or ram, which machine is subjected in use to variable loading and isrequired to operate at different rates; the control system being of the type, herein called the type described, comprising pump control means for varying the swept volume of the pump and prime mover control means, includinga speed responsive governor, for varying the power supplied to the prime mover. By powersupplied to the prime mover, we herein mean rate of energy supply; for example, the rate of supply offuel in the case of an internal comefficient operating conditions for both the pump and the prime mover under varying loadings of the hydraulic machine and with minimum risk of the prime mover being overloaded.

According to this invention we provide in or for a hydraulic transmission apparatus of the kind specified, a control system of the type described, which is characterised by the provision of one operating element, common to both the pump and the prime mover control means and'arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds bustionengine, or the kilowattsin the case where the prime mover is an electric motor.

2. Description of the Prior Art ln hydraulic transmission apparatus of the kind specified, embodying a control system of the type described, if the prime mover and the pump are each to be operated for different loadings of the hydraulic machine, it is necessary for the operator to adjust boththe pump control means and the prime mover control means and to relate these two adjustments to one another if efficiency of operation is to be ensured- For example, if the adjustment is confined to the pump control means so as to reduce the swept volume, the torque demand on the prime mover will be reduced with consequent reduction in fuel or other power utilisation of the prime mover, but in the absence of any positive control on the part of the operator, the prime mover is likely still to be operating near to its maximum speed even when the power demand of the'pump is very small. Under these conditions the prime-mover will be operating in a particularly inefficient manner, e.g., with maximum noise and least economy in fuel consumption in the case of an internal combustion engine. A

If, on the other hand, to meet the particular loading of the hydraulic machine the adjustment of the control means is confined to the prime mover control means so as to reduce the speed thereof, the pump may now be unable to maintain sufficient pressure to meet the load demand without requiring more driving torque than the prime mover is capable of; a condition which can only be overcome by reducing the swept volume of the pump.

Thus ideally if both the pump and the prime mover are to operate efficiently carefully co-ordinated adjustment of both control means is required on the part of the operator, which calls for the exercise of skill and intelligence, so as in particular to avoid so reducing the prime mover power as, in the case of an internal com bustion engine, to cause this to stall.

SUMMARY OF THE INVENTION The present invention has for its object the provision of an improved form of control system of the type described, which readily enables the operator to maintain a predetermined value dependent on the displacement of the operating element, said pump control means permitting the pump swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an' element associated with the prime mover control means for effecting reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latters maximum available torque output.

A control system for an hydraulic transmission according to this invention as above defined in its widest form possesses the following advantages over previously known forms of control system of the type described of which we are aware, namely:

1. The provision of a single operating element so arranged as to effect simultaneous control, i.e., increase of decrease, of both the power supplied to the prime mover and the swept volume of the pump, enables both of these to be controlled efficiently under varying load conditions, without the special skill necessary on the part of the operator when two separate controls have to be carefully manipulated in relation to one another.

2. A control system embodying the present invention also reduces the risk of the prime mover being overloaded by adjustment of the pump capacity to too high a value in relation to the maximum power output of the prime mover.

3. A control system incorporating this invention can readily be so arranged that its effectiveness is not altered by changes in the rating of the particular form of prime mover provided, including for example changes in climate or altitude in the case of an internal combustion engine prime mover or loss of power due to malfunction of the prime mover.

4. Further the system can be readily adapted, by making only minor modifications, for use with a wide range of prime mover/hydraulic transmission apparatus systems.

The present invention is particularly applicable to the control of hydraulic transmission apparatus in which a wide range of speeds of operation on the part of the hydraulic machine are required. For example, in load handling apparatus (such as lifting cranes or winches and in fork trucks) an accurately controlled very slow speed, commonly termed inching, may be needed during the manoevering of the truck to pick up or deposit a load, as well as a considerably higher transportation speed between, e.g., two widely spaced locations in a factory or warehouse complex. iv

The variable swept volume positive displacement pump may be one embodying some form of eccentric control of capacity, for example, a vane type pump in which the swept volume per revolution is adjusted by varying the relative eccentricity of the stator and rotor.

Alternatively the pump may be of the swash plate type embodying a rotatable cluster of pump cylinders, the outer ends of the pistons of which engage with the swash plate, the inclination of which to the axis of rotation of the cluster is adjustable to vary the working stroke of each piston and thus the swept volume of the pump per revolution of the cluster, so as to provide an infinitely variable adjustment between zero and maximum swept volume.

In one form of the invention, the single operating element may be mechanically connected to both the pump control means and the prime mover control means, and the latter may embody a mechanical connection to a power supply operating element including a movable stop element having a lost motion connection to the power supply operating element and, with the latter set to supply maximum power, the mechanical connection acting under the overriding control of the speed responsive governor to engage with a part of the pump control means to inhibit increase in the output of the pump beyond the capacity of the prime mover, the arrangement being such that decrease in prime mover speed at full power supply results in the governor acting to reduce the swept volume of the pump and thus inhibit overriding of the prime mover.

The operating element may comprise a foot pedal so as to provide single pedal control of both the pump and the prime mover.

A mechanically operated system has many practical limitations, for example, restrictions on the permitted relative disposition of pump and prime mover control means if they are both to be controlled mechanically from the one operating element, while mechanically oerated systems inherently may involve severe maintenance problems.

Accordingly, in a preferred form of the control system of this invention, both the pump and the prime mover control means include hydraulic control valves which are simultaneously actuated, e.g., by the operating element, which may comprise a foot pedal, as in the mechanical arrangement above described.

Important features of this preferred form of the invention are:

l. The operation of the pump control valve to control the swept volume of the pump is under the control of both the pump output pressure and also of the governor which, on reduction of prime mover speed with the operating element at maximum setting, operate means controlling the swept volume of the pump to reduce the pump output and thus prevent overloading of the prime mover.

2. The operation of the prime mover control valve to control the power supplied to the prime mover is under the overriding control of a governor controlled valve, the governor being responsive to prime mover speed and operating to control the hydraulic fluid supplied to a power supply operating element to adjust the power supply; e.g., to adjust the fuel supply in the case of an internal combustion engine.

3. The governor controlled valve may include a pressure responsive element and the prime mover control valve may control the amount of hydraulic pressure supplied to the pressure responsive element to control the speed setting, as hereinafter defined. of the governor. The increase in pressure supplied to the pressure responsive element consequent on displacement of the operating element serves to displace the pressure responsive element against the reaction force of the speed responsive governor thereon to increase the speed setting of the governor and hence feed an increased amount of fluid to the power supply operating element to increase the power supply to the prime mover, thus providing for simultaneous increase in engine power supply and speed setting of the governor.

The term speed setting" is used herein toimean the speed at which the governor commences to be operative to inhibit increase in prime mover speed.

4. A hydraulically operated control system lends itself to controlling the operation, when required, of auxiliary machinery so as to take care of the additional power demand on the prime mover, e.g., the hydraulic lifting ram of the forks of a forktruck, namely by providing means for supplying additional pressure to the governor controlled valve to raise the speed setting thereof to give an increase in the power output of the prime mover to provide for the addtional power demand of the auxiliary machinery.

5. A hydraulically operated system readily lends itself to providing a reverse drive from the hydraulic machine which is particularly important where this is a rotary motor for driving a vehicle, namely by the hydraulic control of a change-over valve in the pump control system acting to reverse the operation of the pump control means, e.g., the direction of inclination of the swashplate in the case of a swashplate pump so that the direction of pressure flow from the pump to the motor is now reversed to reverse the direction of operation of the motor.

BRIEF DESCRIPTION OF THE DRAWINGS Two embodiments of the invention are diagrammatically illustrated in the accompanying drawings wherein:

FIG. 1 illustrates a control system of a mechanical form,

FIG. 2 is a hydraulic circuit diagram of a second embodiment of the invention,

FIGS. 3 to 8 inclusive are sectional or part sectional views of certain details of the arrangement depicted in FIG. 2 and FIG. 9 is a hydraulic circuit diagram of a third embodient of the invention, which is a modified form of the second embodiment of the invention and in which the same reference numerals are used in FIG. 9 as in FIG. 2 to refer to corresponding parts.

DESCRIPTION OF THE PREFERRED EMBODIMENTS Referring firstly to FIG. 1 of the drawings the invention is depicted as applied to hydraulic transmission apparatus of the kind specified in which the variable capacity positive displacement pump 20 is of the swashplate type earlier described, the swept volume of which pump can be varied in the known manner by adjusting the angle of the swashplate 21 between the full line position illustrated corresponding to zero pump output and the dashed line position indicated at 21 corresponding to maximum output of the pump.

The output port of the pump is connected to a hydraulic machine in the form of a positive displacement rotary motor 22 and the pump'is driven by a prime mover inthe form of an internal combustion engine, not illustrated but provided with the usual variable fuel supply control 23, the output speed of the engine being under the control of a governor 24. i

The control system for this transmission apparatus comprises pump control means for varying the swept volume of the pump per working stroke and prime mover control means for varying the power supply to the engine, which two control means are, in accordance with this invention, operated bya single operat-- ing element which, as shown, is in the form of a single pivotally mounted foot pedal 25.

The pump control means comprises the aforementioned swashplate 21 which is biassed into the illustrated zero output position by a plunger 26 which is pressurised from the outlet or high pressure port 27 of the pump the displacement of the swashplate be yond the zero position being prevented by stop 28.

The swashplate is displaced from the zero position 21 towards and up to the maximum output position 21 by pumpcontrol plunger 29 forming part of the pump control means, which is engaged by compression spring 30 in engagement with spring stressing rod 31 connected through linkage 32 to the foot pedal 25. Thus as the latter is depressed towards thedashed line position the spring is increasingly stressed toeffect a progressively increased output of the pump by a progressive displacement of the swashplate 21 towards the maximum output position 21 against the biassing load ing of return plunger 26.

The prime mover control means comprise throttle actuating rod 33 carrying arm 34 supporting one end of bar 35, on the other end of which is slidable throttle actuating link 36, which is normally maintained by compression spring 37 in engagement with abutment 38 on the outer end of bar 35. The throttle control rod 33 is supported for sliding movement in the direction of its length and connected at one end to a collar 39 engaged by one end of compression spring 40, the opposite end of which is engaged by spring abutment 41 connected through engine control linkage 42 to the foot pedal 25.

The arrangement is such that when the foot pedal is depressed towards the fully depressed position 25 the spring is increasingly stressed so as, through collar 39, rod 33 and arm 34, to increasingly stress spring 37 and displace the throttle actuating link 36 in a direction from the full line nearly closed position illustrated towards adjustable stop 43 which serves adjustably to limit the maximum opening of the engine throttle or fuel control valve, closing movement of which is simi larly limited by adjustable idling stop 44.

The collar 39 forms part of the governor 24, the weights 45 of which through the bell crank levers 46 act to slide the rod 33 against the pressure of spring 40 in a direction to reduce the fuel supply. Thus increase in the stress in spring 40 by depression of pedal 25 to increase the throttle opening increases the resistance of spring 40 against the action of the governor weights in reducing the throttle opening, i.e., depression of pedal 25 increases the engine speed at which the governor commences to'reduce the throttle opening and hence causes the throttleopening to be increased.

The output port 27 of the pump is connected in the usual way to the inlet port 47 of the motor of which the outlet or low pressure port 48 is connected through the usual low pressure oil reservoir 49 to the low pressure or inlet port 50 of the pump.

From the foregoing it will be understood that with the pedal 25 in the illustrated full line position the swashplate 21 of the pump is in the 'zero output position and the throttle or fuel control valve 23 is in the idling position with the governor weights 45 acting to maintain the engine at its idling speed.

. If the operator now starts to depress the pedal 25, spring 40 will now be stressed so as through spring 37 to displace throttle link 36 towards the open position and at the same time spring 30 will be stressed to effect some displacement of the swashplate 21 away from its zero output position, i.e., into a working position, thus causing the pump 20 to deliver fluid to the motor under the power supplied to the pump from the engine consequent on the increase in engine speed above idling which will now have occurred.

The extent of swashplate movement from the zero position will, however, be limited by the output pressure of the pump acting on plunger 26'so that the angle of inclination of the swashplate 21 will be adjusted until for the particular amount of depression of the pedal 25 the pump output pressure applied to plunger 26 balances the loading of spring 30.

If the operator merely wishes to drive the motor 22 slowly, for example, for the purpose of inching a load powered by the motor such as for the purpose of manoe'uveringa fork truck in which the transmission apparatus is embodied, the pedal 25 is merely depressed by an amount sufiicient to move the pump swashplate from the full line position 21 to provide enough pressure to the motor to enable itto move the load. Such operation may result in a relative] small increase in engine speed above idling with corresponding increase in fuel supply. Insofar as the spring 40 associated with the governor will so far have been stressed to a small extent by a correspondingly small depression of the pedal, the governor despite the quite smallincrease inengine speed will be able to apply pressure to the spring 40 acting in a direction opposite to the pedal pressure, thus limiting the opening movement of the throttle 23, to supply just sufficient fuel to the engine as to maintain the torque demand by the pump at the quite slow motor speed.

' If, on the other hand, the operator requires the load, e.g., a fork truck to be advanced at maximum speed, the pedal 25 will then be depressed to the fullest extent into position 25' thus applying full pressure both to the spring 30 for displacing the swashplate 21 towards the maximum output position 21' and applying full pressure to the governor spring 40. Under these circumstances if the motor 22 was initially at rest the effect of the foregoing pedal operation will initially be to displace the swashplate 21 through quite a small angle out of its illustrated zero position, since such displacement,

with the motor 22 initially at rest, will at once result in the pump 20 starting to deliver fluid at maximum pressure pending acceleration of the motor 22, thus applying full pressure to plunger 26 acting to resist further displacement of the swashplate 21 towards the maximum output position 21'.

At the same time because the maximum pressure has been applied by the pedal to the spring 40, the control rod 33 will have been displaced into a position in which as the engine was still idling, maximum pressure has been applied to the governor spring 40, substantially the whole of which pressure will be transmitted, without significant resistance from the centrifugal reaction on the governor weights, to throttle link 36, so that the throttle 23 will be at once displaced into the full open position, with throttle link 36 now engaging stop 43.

Under the foregoing conditions, as the motor 22 continues to accelerate, its demand on the pump will increase so that the output pressure of the pump will fall somewhat, thus reducing the pressure on plunger 26 and permitting the swashplate 21 advance towards the maximum output position 21' under the sustained pressure of spring 30.

It is usual in such installations as are here considered to avoid an unnecessarily powerful prime mover so that the maximum torque of the prime mover is insufficient to drive the pump both at full pressure and full swept volume. Therefore it is an essential feature of this invention that provision is made for preventing the pump demand on the prime mover reaching a value which the prime mover cannot match.

In accordance with this essential feature provision is made for preventing the pump demand on the engine reaching a value which the engine is unable momentarily to meet with consequent risk of the engine stalling.

The foregoing is taken care of in the arrangement of FIG. 1 by making the control rod 33 of such a length that its free end 51 is adapted as shown at 51' to abut against the swashplate 21 and thus as explained below displace the swashplate in a direction away from the maximum swept volume position.

Under stable conditions any compression of spring 40 will open the throttle valve 23 just sufficiently to enable the prime mover to attain a speed at which the centrifugal force acting on the governor weights produces a force on collar 39 just sufficient to balance the spring load.

In the event of the pump torque requirement exceeding the prime mover torque the prime mover will not attain the speed necessary to balance the load of spring 40 even though rod 33 has moved sufficiently to fully open throttle 23. This out of balance force can be used to move the outer end 51 of rod 33 still further beyond the position shown at 51' compressing the light spring 37 to move swashplate 21 into a position where the stroke of the pump and hence its torque demand is reduced to that available from the prime mover.

It will be appreciated that the above mentioned variation of pump stroke, i.e., swept volume can be a continuing process.

The pressure required of the pump is a function of the load on the motor and the relative speeds of the motor and pump and the stroke of the pump. Thus, if at one point of stability the pressure available is able to accelerate the motor, the pressure after such a period of acceleration will have dropped and hence the torque demand on the prime mover will drop and the prime mover will overspeed and retract rod 33 a little allowing an increase in stroke which in turn will raise the torque demand on the primer mover.

When the swashplate 21 moves eventually to full stroke position or alternatively moves to increase the stroke sufficiently for the system pressure acting on plunger 26 to balance spring 30, further increase in stroke is inhibited as this situation will only have been reached because the prime mover is giving sufficient torque; any further tendency to overspeed will result in movement of rod 33 to the left in the drawing reducing the amount of opening of throttle 23 until prime mover torque is balancing demand, maintaining the prime mover speed at substantially the level demanded by the position of lever 25 and hence the compression of spring 40.

The foregoing co-ordination of the operation of the engine to the power requirements of the motor is effected even though the operator may still depress the pedal 25 to the fullest possible extent.

The load and movement required to operate levers 21 and/or 36 may be too high to enable a satisfactory design based on the above completely mechanical system to be achieved. These difficulties are further enhanced by friction associated with these movements causing unsatisfactory hysteresis effects while the geographical position of the various elements might make mechanical connection difiicult.

These difficulties can be overcome by using power from an available source to make the movements of these levers and control the power by pilot signals.

The available power in practice would normally be either hydraulic or electrical and/or a combination of both could be used as convenient.

To illustrate this a fully hydraulic system of pilot controlled power operated components is now described.

With reference to FIGS. 2 to 8 of the accompanying drawings in which, referring firstly to FIG. 2, the illustrated control system comprises a pump control means including a pump output control valve and a prime mover control means including a prime mover control valve 101 for controlling the speed of the prime mover, which is an internal combustion engine, through a governor valve 102 and hence controlling the power supply to the prime mover. The governor valve 102 is responsive to the output speed of the engine, the fuel supplied to which is determined by the setting of a power supply operating element, in this case a throttle operating unit 104, under the overriding control of valves 101 and 102.

Both the governor 103 and its associated valve 102 are preferably constructed as a single unit (see FlG. 6).

The valves 100, 101 and 102 are a known type of balanced piston valve, with their respective pistons having a pair of axially spaced annular recesses 105, 106 in each case.

The pump output control valve 100 and the engine control valve 10] are both arranged to be simultaneously actuated by a single operating pedal 25 which, as shown, is formed as a bell crank lever and is arranged to displace a pedal bar 107, which through compression springs 108, 109 engaging opposite ends of the bar 107, provides for simultaneous displacement of the pistons of valves 100 and 101. The arrangement is such that the depression of the pedal 25 is capable of increasing the compression of each of these springs and thus increasing by proportional amounts the displacing force which each of these springs exert on these valve pistons. For this purpose the pedal lever 25 is arranged .the latter'engages with the springs 108. 109.

The effect. of so depressing pedal 25 is to allow fluid under pressure from engine driven priming pump 110 which is being continually supplied through line 11 1 to the inlet port of each of thevalves of 100, 101 now to flow intolines 112 and 113 respectively.

Line 112 is connected through reverse valve 114 to one or the other, according to the position of valve 1 14, of lines 115, 116 which respectively lead to forward and reverse direction plungers 117, 118 for displacing against spring loading 119 the swashplate 120 of the pump 121, which is a known form of variable swept volume double acting swashplate pump. The effect of supplying pressure to the forward or reverse swashplate displacing plungers 1 17, 118 is to displace the swashplate 120 from the illustrated full line zero output position in an anti-clockwise or clockwise direction as viewedin FIG. 2, corresponding respectively to .forward or reverse drive of the motor 122 which is a positive displacement rotary motor with the inlet and outlet ports 123, 124 thereof respectively connected to the pump delivery andsuction ports 1 25, 126. l i

. The line 1 13 to which pressure fluid is delivered from the valve 101 is connected through a balancing passage 127 to the end face of the piston of valve 101 opposite to that engaged by,spring 109 and this line 113 is'arranged to supply fluid under pressure through line 113a to one end face of the piston 102a of governor valve 102. The supply is through a two-way valve 128 which serves as later described alternatively to supply fluid to the one end face of governorvalve piston 102a at a pressure higher than that obtaining in'line 113, The effect of increasing the pressure on the one end face of piston 102a is to apply to governor 103 a force acting against centrifugal force and thus increase the speed of rotation of the prime mover at which the governor displaces piston l02a downwardly in FIG. 2 against the pressure in line 113a. That is, the speed setting of the governor is increased.

The governor valve 102 is supplied through line 129 with fluid under pressure from priming pump 110 and the effect of supplying fluid under pressure through line 113 to the governor valve consequent-on the depression of pedal 25 is to displace the piston 102a of valve 102 against the centrifugally derived reaction force of the governor 103 thus, under the overriding control thereof, permitting fluid under pressure to flow from priming pump 110 through line 129 to throttle unit supply line 130.

This line 130 communicates with high pressure port 131 of the cylinder of thejthrottle unit 104 in which slides a throttle displacing plunger 132, the outer end of which is arranged to transmit opening movement to a throttle actuating lever 133 against the loading of a loading spring 133a for this lever.

The plunger 132 is hollow except at its outer end and receives the head 134 of the stem 135 of an anti-stalling piston valve 136, the function of which is to so act, as later described, to reduce the output of the pump when, with the engine throttle fully open, the engine is below its governed speed.

The inner end of plunger 132 is provided with a collar 137 which is arranged to abut the underface of the stem head 134 to effect displacement of the antistalling piston valve 136 as the throttle actuating a plunger 132 is displaced'by more than a certain distance to open the throttle, i.e., to the right in FIG; 2.

The stem 135 at the end thereof, opposite to head 134, is formed with a further head 138 which engages the face of anti-stalling piston valve 136 which is furthest from plunger 132, the opposite face of which piston is subjected to the pressure of return compression spring 139. I

The cylinder of throttle operating unit 104 is provided adjacent piston 136 with an anti-stalling port 140 which, through line 141, leads to return plunger 142, which when pressurised fromline 141 serves through link bar 143 (with which it is held in contact by light spring 142a) to bias the piston of pump control valve 100 in a direction for reducing the pressure suppplied to one or the other of the two swashplate displacing plungers 117, 118. 1

The anti-stalling piston valve 136 is provided with a pair of axially spaced annular ports 136a l36band in the position illustrated corresponding to idling position of the throttle, i.e., with the pedal not depressed, line 141 and associated port 140 is, through piston port 136a, connected to low pressure fluid tank T.

The arrangement is, however, such that the throttle opening movement of plunger 132 is, by its engagement' with stem head 134,'capable of displacing antistalling piston valve 136 to the right in FIG. 2 until port l36b is in communication with port 140.

Port- 136b, through radial passage 136a communicates with the bore of piston 136 which is of a diameter larger than that of stem 135, so that with port 13Gb in communication with anti-stalling port 140, the latter is now pressurized from port 131, subject to the overriding control of the governor valve 102.

' Provision is made for reducing the degree of angularity of the swashplate 120 on increase of pump output pressure beyond a predetermined value, namely by pressure control piston valve 144 and piston 145 having two axially spaced annular ports 145a, l45b which are respectively in communication with the pump ports 125, 126.

.The end faces 1450, 145d of this piston 145 and which are respectively nearest piston ports 145a and 145b, are through lines 115a, 116a, in communication with the swashplate plunger pressurising lines 1 15, 116

which respectively relate to forward and reverse drive of the motor 122.

Thus, for forward drive, for which, under the control of valve 114, lines 115 and 115a are pressurised, the

piston 145 is displaced in a direction for connecting,

through line a the output port 125 of the pump to swashplate angle reducing port 146, while for the reverse direction of pump delivery in which lines 116 and 116a are pressurised so that port 126 which is now the delivery port from the pump, piston will now be displaced in a direction opposite from that just described, thus allowing the fluid at pump delivery pressure to flow through line 126a to swashplate angle reducing port 146.

This port 146 is connected through line 147 to swashplate angle reducing cylinder 148 in which works piston 149 biassed by light spring 149a to maintain its outer end in contact with link bar 143. Piston 149 in like manner to piston 142 is adapted under the pressure in line 147 to displace link bar 143 in a direction to displace the piston of pump control valve 100 against the loading applied thereto from pedal 25.

Because the pressure range in line 147 is the output pressure range of pump 121, very much greater than the pressure range in line 141 derived from the much lower output pressure range of priming pump 110, piston 149 is arranged to act on link bar 143 at a position much nearer to the pivot or fulcrum 143a thereof than is the case with piston 142.

If the pump delivery pressure supplied from valve 144 to swashplate angle reducing piston 149 is great enough to overcome the force of spring 108, the piston of valve 100 will be displaced sufficiently in a direction to close the fluid supply from priming pump 110 through line 112 to the selected swashplate displacing plunger or will even be so displaced as to connect line 112 to the tank T of valve 100. The result in any case is to effect reduction of the angularity of the swashplate under the loading of the appropriate spring 119 thereby reducing the output pressure of the pump 121.

The changeover valve 114 is a simple piston valve which may be operated by reverse drive lever 150 between forward drive and reverse drive positions in which the line supplying the fluid under pressure from pump control valve 100 is alternately connected to lines 115 and 116 leading to forward and reverse drive swashplate displacing plungers 117, 118. in each of these two positions the pressure within the one plunger which is not pressurised is relieved to low pressure tank T through the changeover valve 114 and the piston of this valve also provides an intermediate neutral position in which both of the lines 115 and 116 are simultaneously connected to the low pressure tank.

From the casing of the two-way valve 128 a return pressure line 151 extends from a lift control pilot valve 152, see FIG. 3, the purpose of which is to supply fluid through line 151 and valve 128 at a pressure to line 1 13a greater than that obtaining for the moment in line 113; as for example when pedal 25 is only partly depressed. The effect of supplying fluid at this higher pressure from line 151 is to displace the valve member 128a of valve 128 from its nonnal full line position in which line 113a communicates only with line 113 to the dotted position of FIG. 2 in which line 113a now communicates only with line 151. Thus the higher pressure now supplied to line 1 13a displaces governor valve piston 102a upwardly against the loading of the governor weights to increase the speed at which the governor operates to reduce the throttle opening. There is thus obtained an increased engine speed beyond that resulting from the momentarily obtaining depression of pedal 25.

This feature is particularly advantageous as applied, for example, to fork trucks in which there is an additional demand on the engine when the lifting ram for the fork truck arms is required to be operated for the purpose of raising a load, thus putting on the engine an output demand additional to that required during the transportation of the loaded truck over the floor; it being commonly required to raise a load just picked up by the fork truck arms as the truck is being power advanced along the floor.

Such lift arms or other powered auxiliary devices are normally driven by a hydraulic system entirely separate from that here described but powered from the same engine, hence the necessity to provide the foregoing displacement of governor valve piston 102a to both increase engine speed and power, i.e., fuel supply.

To provide for the foregoing, the lift control pilot valve 152 is provided with a high pressure supply port 153 connected to a high pressure fluid supply, e.g., from priming pump 110.

As shown in FIG. 3, the piston of the pilot valve 152 may be actuated by either or both of two side-by-side pivotally mountedlevers 154, arranged to apply pressure to the end of pilot valve actuating piston 155, which through spring 156 displaces valve piston 157 in a direction for supplying the high pressure fluid from port 153 to outlet port 158 connected to line 151 leading to a valve 128.

As the pressure supplied to line 15] becomes greater than that existing in line 113 the valve member 128a is displaced into the above mentioned dashed line position (FIG. 2) to increase the pressure in line 113a thereby raising the governor speed setting, i.e., the

speed at which the governor 103 operates to reduce the fuel supply to the engine.

When the engine speed is increased as above described, there will be a corresponding increase in the speed of pump 121 and this will tend to increase the speed of motor 122. However, the increase in pump speed will tend to increase the pressure in line 147 leading to cylinder 148 thus having the effect of reducing the angle of the pump swashplate 120 thus compensating for the increase in pump speed and keeping the pump delivery rate substantially constant.

The levers 154 when operated would also in the known way serve to actuate the separate hydraulic system above described.

The hydraulic transmission illustrated in FIG. 2 consisting primarily of pump 121 and motor 122 is of closed circuit configuration which is convenient for a reversingtransmission but which must be maintained full of oil. This is normally achieved by providing a priming pump, which may be the pump or as shown a second priming pump 110a and which is arranged to supply fluid from its delivery port 11% to either line connection 123 or 124, whichever is at the lower pressure, through two-way valve 159 of known form, and in a known manner.

Accordingly opposite ends of the piston 1590 of valve 159 are in communication with lines 160, 161, respectively leading to pump ports 125, 126. According to which of the latter is at the higher pressure, valve piston 159a is moved to connect the priming pump delivery port ll0b to that line 160 or 161 which is at the lower pressure.

If the transmission is used for reversing, e.g., when embodied in a vehicle and is of closed circuit type as illustrated, it is necessary that the pressure signal conveyed in line 147 to piston 149 shall be that from the normal delivery port of pump 121 as determined by reversing valve 114. Thus if pressure in line actuates plunger 1 17 to cause the pump under normal driving conditions to pump fluid from pump port 125 to motor port 124 the same pressure in line 115 will displace valve piston in valve 144 to connect line 125a with port 146 and line 147. If on the other hand the lever 150 is moved to pressurise line 116 instead of line 115 to give reverse drive the piston 145 moves to connect line 1260 to line 147.

As earlier stated the pump control valve 100 and the engine control valve 101 which are arranged to be sirespective springs 108, 109 and pedal bar 107 be embodied in a single unit. The latter may also include valve pistons 142 and 149 with their respective casings together with fulcrum bar 143 and all of these parts may be housed in a common casing so as to form a single unit, the details of which are shown in FIG. 4 and The governor 103, together with the associated valve 102 and valve piston 102a may have the constructional form depicted in FIG. 6, in which these components also form a single unit with the governor, being driven at engine responsive speed through governor shaft 162 on which are mounted the radially movable governor weights l02b which engage with cam plate 163. The latter is supported for sliding movement within housing 164 and through stud 163d engages the adjacent end face of piston 102a so as with increase in engine speed to apply an increasing force to piston 102a in a direction opposite to that obtained from the pressure in line 113a. A light spring is provided for maintaining cam plate 163 in operative engagement with weights l02b when there is no pressure in line 113a.

In FIG. 7 there is depicted in somewhatmore detail the form of the throttle operating unit 104 showing one possible constructional form thereof.

The high pressure lines from-the or each priming pump would be provided with the usual excess pressure relief valves, e.g., valve R in the case of pump 110.

Some of the lines of the above described control sysorifice in the line concerned may suffice at, e.g., the positions 165, shown in FIG. 2 but in other cases, e.g., at some or all of these positions, a special form of damper may be necessary. One preferred form of such a damper is depicted in FIG. 8, consisting of a cylindrical casing 166 in which works a piston 167 loaded against limiting stop 168 by damper spring 169.

The casing 166 provides a smallfchamber 170 in the pressure line in which the damperis provided; the

chamber having inlet and outlet openings l71 and 172 of which the former preferably forms a flow restriction orifice;

. This orifice 171 serves to damp out pulsations in the fluid flow within the line concerned and this is further assisted by the provision of the piston 167, which under sudden pressure increases moves inwardly of the casing 166 against the loading of springl69 so as temporarily to increase the capacity of the chamber 170 and thus relieve any sudden pressure increase.

Pressure relief valves are usually provided associated with and to act between the lines 125a and 126a as a protection against any excess pressure which may arise for any reason.

In operation with the arrangement depicted in FIGS. 2 to 8 and with the pedal 25 in the depicted full-line position of FIG. 2, the piston of pump control valve 100 will be biassed in an upward direction as depicted in FIG. 2 by a pair of light compression springs 142a and 149a associated with pistons 142 and 149 respectively. Thus line 112 through valve 100 is now in communication with the associated low pressure tank T so that neither of the two swashplate displacing plungers 117, 118

are pressurized and the pump swashplate is in the zero output position shown in full in FIG. 2.

The governor valve 102 may be arranged so that the pedal 25 in the'fully up position, maintains a force on spring 109 sufficient to control the governor at a predetermined engine idling speed so that the governor permits a very small fluid flow through line to the throttle valve 104 so as slightly to open the engine throttle and provide for engine idling, any increase in speed above this being precluded by the action of governor'103 in thecompletely closing valve 102.

If now the pedal 25 is depressed slightly the effect is to enable the swashplate to move to obtain a positive pump driving pressure limited to that demanded by the spring 108 at an increase in engine speed limited to that demanded by the spring 109. This is achieved as a result of slight opening of both valves 100 and 101 so that pressure is supplied from the priming pump 1 10 to both lines 112 and 113 through the former to pressurise slightly the particular swashplate plungers selected by the operation of valve 114 and at the same time to pressurise the underside of the piston of the governor valve 102 to increase the speed setting of the governor and hence increase the pressure in the line 130 to the throttle operating unit 104 and thus effect further opening of the throttle.

The effect of so doing, as applied for example to a fork truck, is now to cause the truck to move from rest as soon as the pedal 25 has been depressed sufficiently to obtain enough drive pressure from the pump 121 to produce the necessary driving torque from the motor 122.

Under these circumstances the acceleration of the fork truck or other vehicle will be low and with no further depression of the pedal 25, so that there is only a limited pressure flow through line 113 to the governor valve 102, the governor valve 102 will be prevented from opening further to increase the fuel supply by reason of the overriding action of the governor 103 as the engine speed increases.

Also, if desired, the acceleration of the truck or other vehicle can be stopped or deceleration effected by release of the pedal 25. t

I If, on the other hand,, full acceleration is required the pedal 25 is fully depressed. This, through valve 100, increases the swashplate driving angle towards the maximum consistent with the driving pressure not exceeding the maximum as controlled by the spring 108 in its fully compressed condition and further restricted by the action of the throttle valve 136 which inhibits increased swashplate driving angle if the throttle is in its fully open position indicating that the load on the engine from the pump is excessive and beyond the engine capability. At the same time through valve 101 the maximum engine speed is demanded. As the engine reaches its maximum of speed, plunger 132 will move slightly to the left removing the inhibition of valve 136 to allow the swashplate driving angle to continue to increase up to the maximum driving pressure demand. In the final stabilised position the pump will be at full swashplate drive angle unless the driving pressure required is in excess of the maximum power of the engine and the engine throttle will be controlled to that position which gives enough engine power to satisfy the demand at full engine speed. Thus, when the pedal 25 is fully depressed to fully open the valve 101, this provides for full pressurisation of the underside of the piston 102a of governor valve 102. As the engine speed is as yet quite low, the reaction of the governor weights on piston 1020 is at present toolow to prevent the piston moving into the fully open position to allow full pressure fluid flow through line 130 to the throttle actuat ing plunger 132 which at once moves to the full throttle position, thus ensuring rapid attainment of maximum engine speed concurrently with maximum pump outlet.

For a given pressure in lines 113 and 113a which is controlled by the loading of spring 109, as soon as the reaction force derived from such pressure on governor piston 102a is exceeded by the oppositely acting reaction force from the governor weights, valve 102 will commence to close, thus effecting partial closure of the engine throttle from its previous fully open position. Thus the setting of governor 103, i.e., the speed at which it operates to limit engine speed by partial closure of the throttle, is dependent on the pressure applied through line 1130 to the governor valve piston.

Concurrently with the above described increase in engine speed, as soon as the pump commences to deliver fluid under pressure from the delivery port as determined by the direction of swashplate angularity (controlled by selector valve 114), piston 145 will be biassed in the direction for supplying the pump delivery pressure through line 147 to the underside of piston 149, which is in thrust engagement with fulcrum bar 143, thus through the latter, applying a thrust to the piston of valve 100, tending to displace this from the fully open position against the loading of spring 108, which is a preloaded spring and subject under the foregoing conditions to the pressure of the fully depressed pedal 25.

The characteristics of spring 108 are so selected as to ensure that when the pump delivery pressure in line 147 reaches a predetermined maximum, which may for example, be 3,000 P.S.l., the valve 100 is operated to adjust the pressure supplied through line 112 to the selected swashplate plunger, thus adjusting the swashplate angle, so as to prevent the output pressure of pump 121 exceeding the desired maximum value.

At the same time, with the throttle valve plunger 132 in the fully open position, the flange 137 thereof will, by its engagement with rod head 134, have displaced the associated valve piston 136 against the loading of spring 139 in a direction to supply pressure to line 141, thus pressurising piston 142 upwardly in FIG. 2 to apply further pressure to the piston of valve 100 also acting to adjust the pressure in line 112, thus also having the effect of limiting the output pressure of pump 121.

If with the engine at full throttle opening, the demand of the pump 121 is greater than the power available from the engine so that the speed of the latter commences to fall, the reaction of the governor weights on piston 102a will be less than that derived from the pressure in line 127 assuming that the pedal 25 is still fully depressed and full fuel supply to the engine will thus be maintained to prevent further fall in engine speed under this full power demand condition.

From the foregoing it will be understood that the control system above described is effective in preventing stalling of the engine as a result of excess demand thereon namely by:

a. preventing pump output pressure exceeding a predetermined maximum value,

b. effecting reduction of pump output when the fuel supply control exceeds a certain maximum value,

c. maintaining full fuel supply when the engine speed has fallen below that demanded by the operation of pedal 25.

' Deceleration is effected by partially or fully releasing the pedal 25. If the pedal is partially released the engine speed is reduced and this, with the vehicle moving, will cause the motor 122 to drive the pump 121 and thus there will be a severe reduction in pressure acting through the line 147 on the piston 149. Thus any load on the spring 108 will move the piston of the valve to cause the swashplate angle to increase. As mentioned above under these conditions the motor 122 will be driving the pump 121 and hence driving the engine at a speed higher than the governor setting which will therefore reduce the fuel supply to the minimum and provide a braking effort on the vehicle comparable with the torque required to motor the engine.

If the pedal 25 is fully released, then the load on the spring 108, may, if hydraulic braking is required, be arranged to reduce to zero in which case the springs 142a and 149a will cause the piston of the valve 100 to rise and reduce the swashplate angle to zero, or near to zero. Under these conditions there will be a high hydraulic resistance against the pumping action of the motor which will provide a large hydraulic braking effect limited by any pressure release valves associated with the lines and 126.

As already mentioned, during rapid decleration of the vehicle so that the motor 122 is now acting as a pump, the deceleration torque applied to the motor will produce a fluid flow through the line which will assist in braking, the pressure of which will, however, be the reverse of that obtaining under forward drive conditions but since there has been no reversal of the selector valve 114, reverse valve 144 will not have reversed its position and there will consequently be no transmission of pressure from the motor to line 114 which without change in position of the selector valve 114 is still in communication with line 112.

Thus, under the foregoing conditions the return pressure developed within the motor will not be transmitted back to the controlling valves.

If desired the motor 122 may be of the variable swept volume type. Such a provision enables the swept volume of the motor to be reduced for the purpose of obtaining a higher motor speed, with corresponding torque reduction, beyond the maximum speed obtained with full fuel supply and full swept volume of the pump; e.g., for fast vehicle running where the torque demand on the motor is below maximum.

In the embodiment of the invention described with reference to FIGS. 2 to 8, as described above, if the pedal 25 is considerably depressed so that the vehicle achieves a relatively high speed and then the pedal 25 is partially released, the effect will be for the control system to attempt to reduce the engine speed whilst the motor, 122, is continued to be driv4en by the vehicle on over-run and hence will drive the engine through the pump 121.

Under these circumstances pressure in the line 147 to the swashplate angle reducing cylinder 148 falls to a low value and hence any force still applied to the spring 108 will cause the swashplate of the pump 121 to move towards its greatest angle, if it is not already at this angle. The vehicle will therefore be braked by the effort of the motorto drive the engine.

It is however possible to obtain a much larger braking force when the vehicle. is moving by reducing the swashplate angle towards zero. Thisaction chokes the motorcircuit and this putsia back pressure onto the motor and although suchbraking effort is in practice limited by release valves in the lines 125 and 126 as described hereinbefore can nevertheless become very severe.

In order to achieve a more controlled form of hydraulic braking the embodiment described with reference to FIGS. 2 to 8 may be modified as shown in FIG. 9.

Turning now to FIG. 9 the pedal 25 and the pedal bar 107 are produced as a single component and in this embodiment the prime mover control valve 101 and associated element are disposed so that they are behind the pump control valve 100 and associated elements and so are not shown in FIG. 9. The pedal bar 107 is arranged to act on the valves 100 and 101 in the same way as described with reference to FIG. 2. The pedal 25 is provided with an extension part 200 on the opposite side of its pivot to the pedal bar 107. The extension part 200 is arranged to operate on a coil compression spring 201 which acts on a nose element 202 connected through a lost motion device 203 to a spring abutment element 204. The nose element 202 acts on an etension part 205 of the link'bar 143. Also acting on the extension part 205 is a second piston 210 in a cylinder 211 similar to the piston 149 and cylinder 148 arranged to pivot the bar 143 in the opposite direction to the'directions in which it is pivoted by the piston 149.- The pressure controlled piston valve 144 in this embodiment takes theform of a two port two positioned valve 220 which is again under the control'of the forward and reverse drive lever 150 and which again permits the normal drive pressure line of the main hydraulic circuit to be connected to the piston 149, through the line 147 as in the emboidment described with reference to FIG. 2 and in addition permits the return line of the main hydraulic circuitto be connected to a cylinder 211 of the piston 210 through the line 221.

The pedal 25 is acted upon by a coil tension spring 222 which is of such rate as to return the pedal 25 to its fully released position against any resistance applied thereto by the spring 201.

it is envisaged that the spring 108 will probably be affected by the full movement of the lever 25 whilst the spring 201 will be restricted so that it is only operated as the pedal 25 approaches its fully released position. At some intermediate position of the pedal 25 the springs 108 and 201 will balance each other so that the system is set for equal pressure to exist in the main drive and return lines of the main hydraulic circuit. lf the pressure in the main drive line should predominate then the pressure therein will be fed by the line 147 to the cylinder 148 to cause the piston 149 therein to move to pivot the link bar 143 clockwise in FIG. 9 to move the piston of the valve 100 to decrease the swashplate angle and so reduce the pressure in the main drive line.

If, on the other hand, the pressure in the return line predominates then the pressure therein fed by the line 221 to the cylinder 211 would cause the piston 210 therein to pivot the-bar 143anti-clockwise in FIG. 9 to cause the piston of the valve 100 to move to increase the swashplate angle and so reduce the pressure in the return line.

Depression of the pedal 25 beyond the point of balance between the springs 20] and 108 will cause drive to occur as described with reference to FlGS. 2 to 8.

Conversely, movement of the pedal 25 upwardly of the point of spring balance will cause a predominance of force from the spring 201 and will cause the link bar 143 to pivot clockwise in FIG. 9 to move the piston of the valve upwardly and thus to decrease the swashplate angle. lf thevehicle is moving this reduction of swashplate angle will cause a braking pressure to build up in the return line of the main circuit which pressure will be transmitted by the line 221 to the cylinder 211 and hence the pressure will be limited to that pressure which, acting on the piston 210, produces a force which through the link bar 143 and extension part 205 thereof is capable of balancing the force on the spring 201.

I claim:

1. A control system in hydraulic transmission apparatus comprising a prime mover, a variable swept volume positive displacement pump driven by said prime mover, a hydraulic machine operated by said pump, said machine being subjected, in use, to variable loading and being required to operate at different rates; the control system comprising, pump control means for varying the swept volume of the pump, prime mover control means for varying the power supplied to the prime mover, one operating element common to both the pump and prime mover. control means and arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds a predetermined value dependent on the displacement of the operating element, said pump control means permitting the swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means including a speed responsive governor and being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an element associated with the prime mover control means for effecting reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latters maximum available torque output.

2. A control system according to claim 1 wherein the pump control means includes resilient biasing means which is increasingly stressed to effect an increase in swept volume on displacement of the operating element from the idling position, which biassing means is also capable of being further stressed under pressure derived from the pump output in such a manner as to allow the pump control means to reduce the swept volume even though the operating element may already be exerting its maximum pressure on the biassing means.

3. A control system according to claim 1 wherein the prime mover control means comprises a further resilient biassing means which is increasingly stressed to increase the speed setting of the governor on displacement of the operating element from the idling position to increase the power supplied.

4. A control system according to claim 1 wherein the single operating element is mechanically connected to both the pump control means and the prime mover control means and the latter embodies a mechanical connection, to a power supply operating element, including a movable stop element having a lost motion connection to the power supply operating element and, with the latter set to supply maximum power, the mechanical connection acting under the overriding control of the speed responsive governor to engage with a part of the pump control means to inhibit increase in the output of the pump beyond the capacity of the prime mover, the arrangement being such that decrease in prime mover speed at full power supply results in the governor acting to reduce the swept volume of the pump and thus inhibit overloading of the prime mover.

5. A control system according to claim 1 wherein both the pump and the prime mover control means each include a hydraulic control valve which are simultaneously actuated by said operating element, operation of the pump control valve to control the swept volume of the pump being under the control of both the pump output pressure and also of the governor, which on reduction of prime mover speed with the power supply operating element at maximum setting, operates means controlling the swept volume of the pump to reduce the pump output and thus inhibit overloading of the prime mover.

6. A control system according to claim 5 wherein said control system includes means responsive to said pump output pressure and operative on the pump control valve to cause the valve to operate to reduce the swept volume of the pump when the pump output pressure exceeds said predetermined value.

7. A control system according to claim 5 wherein operation of the prime mover control valve to control the power supplied to the prime mover is under the overriding control of a governor controlled valve, the

governor being responsive to prime mover speed and operating to control the hydraulic fluid supplied to the power supply operating element to adjust the power supply.

8. A control system according to claim 7 wherein the governor controlled valve includes a pressure responsive element, the prime mover control valve controls the amount of hydraulic pressure applied to said pressure responsive element and wherein increase in pressure supplied to the pressure responsive element displaces said pressure responsive element against the reaction force of the speed responsive governor thereon to increase the speed setting of the governor to feed an increased amount of fluid to the power supply operating element, thereby to incease the power supply to the prime mover.

9. A control system according to claim 1 including means raised to effect increase in the swept volume of the pump when the return pressure thereof exceeds a pedetermined value dependent on the displacement of the operating element.

10. A control system according to claim 1 wherein both the pump and the prime mover control means each include a hydraulic control valve which are simultaneously actuated by said operating clement, operation of the pump control valve to control the swept volume of the pump being under the control of both the pump output pressure and also of the governor, which on reduction of prime mover speed with the power supply operating element at maximum setting, operates means controlling the swept volume of the pump to reduce the pump output and thus inhibit overloading of the prime mover, and the control system includes means responsive to said pump return pressure and operative on the pump control valve to cause the valve to operate to increase the swept volume of the pump when said pump return pressure exceeds said predetermined value. 

1. A control system in hydraulic transmission apparatus comprising a prime mover, a variable swept volume positive displacement pump driven by said prime mover, a hydraulic machine operated by said pump, said machine being subjected, in use, to variable loading and being required to operate at different rates; the control system comprising, pump control means for varying the swept volume of the pump, prime mover control means for varying the power supplied to the prime mover, one operating element common to both the pump and prime mover control means and arranged to effect simultaneous control of both the swept volume of the pump and of the power supplied to the prime mover, said pump control means being arranged to effect reduction in the swept volume of the pump when the output pressure thereof exceeds a predetermined value dependent on the displacement of the operating element, said pump control means permitting the swept volume to be so reduced despite displacement of the operating element to, or towards, the position calling for maximum swept volume, said prime mover control means including a speed responsive governor and being arranged to control the power supplied to the prime mover by controlling the speed setting of the governor and means responsive to an element associated with the prime mover control means for effecting reduction in the swept volume of the pump when the torque load from the pump on the prime mover commences to exceed the latter''s maximum available torque output.
 2. A control system according to claim 1 wherein the pump control means includes resilient biasing means which is increasingly stressed to effect an increase in swept volume on displacement of the operating element from the idling position, which biassing means is also capable of being further stressed under pressure derived from the pump output in such a manner as to allow the pump control means to reduce the swept volume even though the operating element may already be exerting its maximum pressure on the biassing means.
 3. A control system according to claim 1 wherein the prime mover control means comprises a further resilient biassing means which is increasingly stressed to increase the speed setting of the governor on displacement of the operating element from the idling position to increase the power supplied.
 4. A control system according to claim 1 wherein the single operating element is mechanically connected to both the pump control means and the prime mover control means and the latter embodies a mechanical connection, to a power supply operating element, including a movable stop element having a lost motion connection to the power supply operating element and, with the latter set to supply maximum power, the mechanical connection acting under the overriding control of the speed responsive governor to engage with a part of the pump control means to inhibit increase in the output of the pump beyond the capacity of the prime mover, the arrangement being such that decrease in prime mover speed at full power supply results in the governor acting to reduce the swept volume of the pump and thus inhibit overloading of the prime mover.
 5. A control system according to claim 1 wherein both the pump and the prime mover control means each include a hydraulic control valve which are simultaneously acTuated by said operating element, operation of the pump control valve to control the swept volume of the pump being under the control of both the pump output pressure and also of the governor, which on reduction of prime mover speed with the power supply operating element at maximum setting, operates means controlling the swept volume of the pump to reduce the pump output and thus inhibit overloading of the prime mover.
 6. A control system according to claim 5 wherein said control system includes means responsive to said pump output pressure and operative on the pump control valve to cause the valve to operate to reduce the swept volume of the pump when the pump output pressure exceeds said predetermined value.
 7. A control system according to claim 5 wherein operation of the prime mover control valve to control the power supplied to the prime mover is under the overriding control of a governor controlled valve, the governor being responsive to prime mover speed and operating to control the hydraulic fluid supplied to the power supply operating element to adjust the power supply.
 8. A control system according to claim 7 wherein the governor controlled valve includes a pressure responsive element, the prime mover control valve controls the amount of hydraulic pressure applied to said pressure responsive element and wherein increase in pressure supplied to the pressure responsive element displaces said pressure responsive element against the reaction force of the speed responsive governor thereon to increase the speed setting of the governor to feed an increased amount of fluid to the power supply operating element, thereby to incease the power supply to the prime mover.
 9. A control system according to claim 1 including means raised to effect increase in the swept volume of the pump when the return pressure thereof exceeds a pedetermined value dependent on the displacement of the operating element.
 10. A control system according to claim 1 wherein both the pump and the prime mover control means each include a hydraulic control valve which are simultaneously actuated by said operating element, operation of the pump control valve to control the swept volume of the pump being under the control of both the pump output pressure and also of the governor, which on reduction of prime mover speed with the power supply operating element at maximum setting, operates means controlling the swept volume of the pump to reduce the pump output and thus inhibit overloading of the prime mover, and the control system includes means responsive to said pump return pressure and operative on the pump control valve to cause the valve to operate to increase the swept volume of the pump when said pump return pressure exceeds said predetermined value. 